JP4811153B2 - Control device for automatic transmission - Google Patents

Description

  The present invention relates to a control device for an automatic transmission having a pressure adjusting device that adjusts hydraulic oil discharged from an oil pump to line hydraulic pressure based on a hydraulic pressure command value, and in particular, for obtaining necessary line hydraulic pressure appropriately. It is about technology.

  Equipped with a pressure regulator that regulates the hydraulic oil discharged from the oil pump to the line oil pressure based on the oil pressure command value, and the oil pressure command value based on the required line oil pressure from the pre-stored relationship between the oil pressure command value and the line oil pressure A control device for an automatic transmission that outputs the motor is well known. For example, the pressure regulator is composed of a regulator valve and a linear solenoid valve, and an output hydraulic pressure of the linear solenoid valve corresponding to a hydraulic pressure command value, that is, a control current value for obtaining a required line hydraulic pressure acts on the regulator valve. As a result, the line hydraulic pressure is electrically controlled.

  In such a pressure regulating device, the relationship (map) stored in advance between the hydraulic pressure command value and the line hydraulic pressure varies or changes due to individual differences or changes with time. Therefore, in Patent Document 1, when the actual line oil pressure adjusted by the pressure adjusting device is detected, and the difference between the stored map and the latest map of the oil pressure command value and the actual line oil pressure is large, There has been proposed an automatic transmission control device that learns and corrects the latest map so that an accurate map is obtained.

JP 2005-163934 A

  By the way, the actual line oil pressure may include a rising amount, which is an increase in oil pressure that varies with changes in the rotation speed of the oil pump, with the rotation speed of the oil pump as an influencing factor. Then, even if the oil pressure command value is the same, the actual line oil pressure varies depending on the rotation speed of the oil pump. Therefore, if the map is learned and corrected without taking the rising amount into consideration as in Patent Document 1, it is accurate. There was a possibility that a correct map could not be obtained. That is, there is a problem that the learning accuracy of the map deteriorates.

  The automatic transmission has a driving pulley, a driven pulley, and a belt wound around both pulleys. The transmission ratio is controlled by the flow rate of hydraulic oil to the driving pulley and the driven pulley is In the case of a continuously variable transmission in which the belt clamping pressure is controlled by the hydraulic pressure of the working hydraulic fluid, the hydraulic fluid to the driving pulley is based on the differential pressure between the line hydraulic pressure and the hydraulic pressure of the hydraulic fluid acting on the driving pulley. It is sometimes necessary to obtain an estimated value of the line oil pressure when performing the shift control by calculating the flow rate of the engine. In such a case, it is conceivable to calculate the estimated value of the line oil pressure using a learning-corrected map. However, as described above, if the rising amount is not taken into account, there is a problem that the estimation accuracy of the line oil pressure deteriorates. there were.

  The present invention has been made in the background of the above circumstances, and its object is to improve the learning accuracy of the relationship between the hydraulic pressure command value and the line hydraulic pressure in the automatic transmission control device. .

  The gist of the invention according to claim 1 for achieving the object is as follows: (a) a pressure adjusting device that adjusts hydraulic oil discharged from an oil pump to line hydraulic pressure based on a hydraulic pressure command value; A line hydraulic pressure control means for outputting a hydraulic pressure command value for obtaining the required line hydraulic pressure based on the required line hydraulic pressure from a previously stored relationship between the value and the line hydraulic pressure, and the pressure regulating device A control device for an automatic transmission comprising actual line oil pressure detection means for detecting actual line oil pressure, and learning correction means for learning and correcting the relationship based on the actual line oil pressure detected by the actual line oil pressure detection means. And (b) further comprising a hydraulic pressure increase calculating means for calculating a hydraulic pressure increase of the line hydraulic pressure based on the rotation speed of the oil pump and using the rotation speed of the oil pump as an influencing factor , (C) the learning correction means is to learn more correct the relationship based on the hydraulic rise calculated by the pressure rise calculating means.

  In this way, based on the actual line oil pressure detected by the actual line oil pressure detection means and the oil pressure increase of the line oil pressure having the rotational speed of the oil pump calculated by the oil pressure increase calculation means as an influencing factor, The learning correction means learns and corrects the previously stored relationship between the hydraulic pressure command value and the line hydraulic pressure, so that an appropriate relationship between the hydraulic pressure command value and the line hydraulic pressure can be obtained regardless of the increase in the hydraulic pressure of the line hydraulic pressure. The learning accuracy of the relationship between the oil pressure command value and the line oil pressure is improved. As a result, for example, the oil pressure command value can be set so that the line oil pressure is as small as possible within a range where the necessary line oil pressure can be obtained, and the line oil pressure is set higher than necessary. Compared with the case where the hydraulic pressure command value is set to, the loss of the oil pump is suppressed and the fuel consumption is improved.

  The invention according to claim 2 is the control device for an automatic transmission according to claim 1, wherein the automatic transmission includes a driving pulley, a driven pulley, and a belt wound around both pulleys. A continuously variable transmission in which a transmission ratio is controlled by hydraulic fluid acting on the driving pulley and a belt clamping pressure is controlled by hydraulic fluid acting on the driven pulley; learning by the learning correction means From the corrected relationship, an estimated value of the line oil pressure is calculated based on the oil pressure command value, and the line oil pressure estimated value calculating means for correcting the estimated value of the line oil pressure based on the hydraulic pressure increase is further provided. The continuously variable transmission is controlled based on the estimated value of the line hydraulic pressure. By doing so, the estimation accuracy of the line hydraulic pressure is improved. This improves the control accuracy of the continuously variable transmission. For example, when performing the shift control by calculating the flow rate of the hydraulic oil to the drive pulley based on the differential pressure between the line hydraulic pressure and the hydraulic oil pressure acting on the drive pulley, the required line hydraulic pressure The estimated value is obtained with high accuracy, and the shift control accuracy is improved.

  Here, preferably, the automatic transmission is disposed in a power transmission path between the driving power source and the driving wheel, and the rotating elements of the plurality of sets of planetary gear devices are configured to have hydraulic friction. A plurality of gear stages (shift stages) are selectively achieved by being selectively connected by the engagement device, for example, having four forward stages, five forward stages, six forward stages, and more. Various planetary gear type multi-stage transmissions such as the above, a so-called belt type non-transmissible belt in which a transmission belt functioning as a power transmission member is wound around a pair of variable pulleys having variable effective diameters and the gear ratio is continuously changed steplessly. Consists of a step transmission and the like.

  Preferably, in the normal transmission control of the continuously variable transmission, for example, a primary transmission-side hydraulic cylinder is obtained so that a target transmission ratio is obtained according to a predetermined transmission condition and the actual transmission ratio becomes the target transmission ratio. By changing the primary pulley groove width by controlling the flow rate of hydraulic oil to the vehicle, the gear ratio is feedback-controlled, or the input side (drive source side) according to the vehicle speed, output rotational speed (drive wheel side rotational speed), etc. ), And by changing the primary pulley groove width by controlling the flow rate of hydraulic oil to the primary pulley side hydraulic cylinder so that the actual input rotational speed becomes the target rotational speed, Various modes such as feedback control can be adopted.

  The above-mentioned predetermined speed change conditions are set by a map or an arithmetic expression using, for example, driving conditions such as a driver output request amount (acceleration request amount) such as an accelerator operation amount and a vehicle speed (corresponding to an output rotation speed) as parameters Is done.

  Preferably, an engine that is an internal combustion engine such as a gasoline engine or a diesel engine is widely used as the driving power source. Furthermore, an electric motor or the like may be used in addition to the engine as an auxiliary driving power source. Alternatively, only an electric motor may be used as a driving power source.

  Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings.

  FIG. 1 is a skeleton diagram illustrating the configuration of a vehicle drive device 10 to which the present invention is applied. This vehicle drive device 10 is a horizontal automatic transmission, which is suitably employed in an FF (front engine / front drive) type vehicle, and includes an engine 12 as a driving power source. The output of the engine 12 composed of an internal combustion engine is the crankshaft of the engine 12, the torque converter 14 as a fluid transmission device, the forward / reverse switching device 16, the belt type continuously variable transmission (CVT) 18, the reduction gear. It is transmitted to the differential gear device 22 via the device 20 and distributed to the left and right drive wheels 24L, 24R.

  The torque converter 14 includes a pump impeller 14p connected to the crankshaft of the engine 12 and a turbine impeller 14t connected to the forward / reverse switching device 16 via a turbine shaft 34 corresponding to an output side member of the torque converter 14. And power transmission is performed via a fluid. A lock-up clutch 26 is provided between the pump impeller 14p and the turbine impeller 14t, and a lock-up control valve (L not shown) in the hydraulic control circuit 100 (see FIGS. 2 and 3) is provided. The hydraulic pressure supply to the engagement side oil chamber and the release side oil chamber is switched by the (C / C control valve) or the like, thereby being engaged or released. The turbine impeller 14t is rotated integrally. The pump impeller 14p has a hydraulic pressure for controlling the transmission of the continuously variable transmission 18, generating a belt clamping pressure, controlling the engagement release of the lockup clutch 26, or supplying lubricating oil to each part. Is coupled to a mechanical oil pump 28 that is generated by being driven to rotate by the engine 12.

  The forward / reverse switching device 16 is mainly composed of a double pinion type planetary gear device, the turbine shaft 34 of the torque converter 14 is integrally connected to the sun gear 16s, and the input shaft 36 of the continuously variable transmission 18 is a carrier. The carrier 16c and the sun gear 16s are selectively connected via the forward clutch C1, and the ring gear 16r is selectively fixed to the housing via the reverse brake B1. It has become. The forward clutch C1 and the reverse brake B1 correspond to an intermittent device, both of which are hydraulic friction engagement devices that are frictionally engaged by a hydraulic cylinder.

  When the forward clutch C1 is engaged and the reverse brake B1 is released, the forward / reverse switching device 16 is brought into an integral rotation state, whereby the turbine shaft 34 is directly connected to the input shaft 36, and the forward power The transmission path is established (achieved), and the driving force in the forward direction is transmitted to the continuously variable transmission 18 side. When the reverse brake B1 is engaged and the forward clutch C1 is released, the forward / reverse switching device 16 establishes (achieves) the reverse power transmission path, and the input shaft 36 is connected to the turbine shaft 34. On the other hand, it is rotated in the opposite direction, and the driving force in the reverse direction is transmitted to the continuously variable transmission 18 side. Further, when both the forward clutch C1 and the reverse brake B1 are released, the forward / reverse switching device 16 becomes neutral (interrupted state) for interrupting power transmission.

  The continuously variable transmission 18 includes an input-side variable pulley (drive pulley, primary pulley) 42 having a variable effective diameter, which is an input-side member provided on the input shaft 36, and an output-side member provided on the output shaft 44. An output-side variable pulley (driven pulley, secondary pulley) 46 having a variable effective diameter and a transmission belt 48 wound around the variable pulleys 42 and 46 are provided. Power is transmitted through frictional force with the belt 48.

The variable pulleys 42 and 46 are fixed rotation bodies 42 a and 46 a fixed to the input shaft 36 and the output shaft 44, respectively, and are not rotatable relative to the input shaft 36 and the output shaft 44 and are movable in the axial direction. Provided movable rotating bodies 42b and 46b, and an input side hydraulic cylinder (primary pulley side hydraulic cylinder) 42c and an output side hydraulic cylinder (secondary pulley side hydraulic cylinder) 46c that apply thrust to change the V groove width therebetween. The hydraulic fluid supply / discharge flow rate to the input side hydraulic cylinder 42c is controlled by the hydraulic control circuit 100, so that the V-groove widths of both variable pulleys 42 and 46 are changed to change the transmission belt. changed consuming diameter of 48 (effective diameter), the speed ratio gamma (= input shaft speed _{N iN} / output shaft speed _{N OUT)} continuous It is changed to. Further, the hydraulic pressure of the output side hydraulic cylinder 46c (belt clamping pressure Pd) is regulated by the hydraulic control circuit 100, whereby the belt clamping pressure is controlled so that the transmission belt 48 does not slip. As a result of such control, the hydraulic pressure (shift control pressure Pin) of the input side hydraulic cylinder 42c is generated.

  FIG. 2 is a block diagram for explaining a main part of a control system provided in the vehicle for controlling the vehicle drive device 10 of FIG. The electronic control unit 50 includes, for example, a so-called microcomputer having a CPU, a RAM, a ROM, an input / output interface, and the like. The CPU uses a temporary storage function of the RAM and follows a program stored in the ROM in advance. By performing signal processing, output control of the engine 12, shift control of the continuously variable transmission 18, belt clamping pressure control, torque capacity control of the lockup clutch 26, and the like are executed. This is divided into control and hydraulic control for the continuously variable transmission 18 and the lockup clutch 26.

The electronic control unit 50, representing the crankshaft rotation speed corresponding to the engine rotational speed crankshaft detected by the sensor 52 rotation angle (position) A _{CR (°)} and the rotational speed of the engine 12 (engine rotational speed) N _E Signal, a signal representing the rotational speed (turbine rotational speed) _NT of the turbine shaft 34 detected by the turbine rotational speed sensor 54, and an input that is the input rotational speed of the continuously variable transmission 18 detected by the input shaft rotational speed sensor 56. rotational speed (input shaft rotational speed) signal representing the N _iN of the shaft 36, a vehicle speed sensor (output shaft rotation speed sensor) 58 by the rotational speed (output of the output shaft 44 is the output rotational speed of the continuously variable transmission 18 detected axis rotation speed) N _OUT ie vehicle speed signal representing a vehicle speed V corresponding to the output shaft speed N _OUT, en detected by the throttle sensor 60 Intake pipe 32 throttle valve opening signal representing the throttle valve opening theta _TH of the electronic throttle valve 30 provided in (see FIG. 1) of the emission 12, the cooling water temperature T _W of the engine 12 detected by a coolant temperature sensor 62 signals representative signal representative of the oil temperature T _CVT of a hydraulic circuit of the continuously variable such as transmission 18 detected by the CVT oil temperature sensor 64, the accelerator opening is an operation amount of the accelerator pedal 68 detected by the accelerator opening sensor 66 an accelerator opening signal representative of the acc, brake operation signal indicating whether B _ON operation of the foot brake is a service brake, which is detected by a foot brake switch 70, a lever position (operation of the shift lever 74 detected by a lever position sensor 72 Position) An operation position signal representing _PSH is supplied.

Further, the electronic control device 50 receives an engine output control command signal S _E for controlling the output of the engine 12, for example, a throttle signal for driving a throttle actuator 76 for controlling the opening / closing of the electronic throttle valve 30, and a fuel injection device 78. An injection signal for controlling the amount of fuel injected from the engine, an ignition timing signal for controlling the ignition timing of the engine 12 by the ignition device 80, and the like are output. Further, a command for driving the solenoid valve DS1 and the solenoid valve DS2 for controlling the flow of hydraulic fluid to the shift control command signal S _T, for example, the input-side hydraulic cylinder 42c for changing the speed ratio γ of the continuously variable transmission 18 signal, a command signal for driving the squeezing force control command signal S _B for example, a linear solenoid valve pressure belt clamping pressure Pd adjusted SLS for aligning clamping pressure of the transmission belt 48, the line for which control the line pressure P _L such command signal for driving the hydraulic control command signal S _PL example, a linear solenoid valve pressure line pressure P _L tone SLT is output to the hydraulic control circuit 100.

  The shift lever 74 is arranged, for example, in the vicinity of the driver's seat and is sequentially positioned in five lever positions “P”, “R”, “N”, “D”, and “L” (see FIG. 3). Any one of them is manually operated.

  The “P” position (range) releases the power transmission path of the vehicle drive device 10, that is, a neutral state (neutral state) where the power transmission of the vehicle drive device 10 is interrupted, and is mechanically output by the mechanical parking mechanism. The parking position (position) for preventing (locking) the rotation of 44, the “R” position is the reverse traveling position (position) for reversely rotating the output shaft 44, and the “N” position. Is a neutral position (position) for setting the neutral state in which the power transmission of the vehicle drive device 10 is interrupted, and the “D” position establishes an automatic transmission mode within a transmission range that allows the transmission of the continuously variable transmission 18. This is a forward travel position (position) that allows automatic shift control to be executed, and the “L” position is operated by a strong engine brake. An engine brake position (position). Thus, the “P” position and the “N” position are non-traveling positions that are selected when the vehicle is not traveling, and the “R” position, the “D” position, and the “L” position are when the vehicle is traveling. This is the selected driving position.

  FIG. 3 shows the main part of the hydraulic control circuit 100 relating to the belt clamping pressure control, the transmission gear ratio control of the continuously variable transmission 18, and the engagement hydraulic pressure control of the forward clutch C1 or the reverse brake B1 accompanying the operation of the shift lever 74. FIG. In FIG. 3, the hydraulic control circuit 100 continuously changes the clamping pressure control valve 110 that regulates the belt clamping pressure Pd that is the hydraulic pressure of the output hydraulic cylinder 46 c so that the transmission belt 48 does not slip, and the speed ratio γ continuously changes. The transmission ratio control valve UP114 and the transmission ratio control valve DN116 for controlling the flow rate of hydraulic fluid to the input side hydraulic cylinder 42c so that the ratio between the transmission control pressure Pin and the belt clamping pressure Pd is set in a predetermined relationship. A manual valve 120 or the like that mechanically switches the oil path according to the operation of the shift lever 74 is provided so that the thrust ratio control valve 118, the forward clutch C1 and the reverse brake B1 are engaged or released.

Line pressure P _L as source pressure working oil pressure output from the mechanical oil pump 28 (see FIG. 1) which is rotated by the engine 12 (generator), for example, a primary regulator valve (line oil pressure regulating valve of the relief type ) 122 is adjusted to a value corresponding to the engine load or the like based on the control hydraulic pressure P _SLT which is the output hydraulic pressure of the linear solenoid valve SLT.

More specifically, the primary regulator valve 122 is provided so as to be movable in the axial direction, thereby opening and closing the input port 122i and discharging the hydraulic pressure generated from the oil pump 28 to the intake oil passage 124 via the output port 122t. The spool valve element 122a, the spring 122b as an urging means for urging the spool valve element 122a in the valve closing direction, and the spring 122b for accommodating the spool valve element 122a and applying a thrust force in the valve closing direction to the spool valve element 122a. the control oil pressure and oil chamber 122c for receiving the P _SLT, and an oil chamber 122d that receives the hydraulic pressure generated from the oil pump 28 to apply a thrust force in the valve opening direction to the spool valve element 122a to.

In the primary regulator valve 122 configured as described above, and the biasing force _{F S} of the spring 122b, the pressure receiving area of the control oil pressure _{P SLT} in the oil chamber 122c a, the pressure receiving area difference of the line pressure _{P L} in the oil chamber 122d b Then, it will be in an equilibrium state by following Formula (1).
P _L × b = P _SLT × a + F _S (1)
Therefore, the line pressure _{P L} is represented by the following formula (2), is proportional to the control pressure _{P SLT.}
P _L = P _SLT × (a / b) + F _S / b (2)

Thus, the primary regulator valve 122 and the linear solenoid valve SLT, a line oil pressure control command signal S _PL pressure regulating pressure regulating hydraulic oil to the line pressure P _L to be discharged from the oil pump 28 on the basis of as the hydraulic pressure command value Functions as a device.

Modulator pressure P _M, as well is used as the basic pressure of the control oil pressure P _SLS is the output hydraulic pressure of the control pressure P _SLT and the linear solenoid valve SLS, by the electronic control unit 50 by the output hydraulic pressure of the solenoid valve DS1 that is duty-controlled a used as the basic pressure of a certain control oil pressure P _DS1 and the control pressure P _DS2 is the output hydraulic pressure of the solenoid valve DS2, the modulator valve 126 to line pressure P _L as source pressure adapted to be pressure regulated to a constant pressure ing.

Output hydraulic pressure _{P LM2} is adapted to line pressure _{P L} to be pressure regulated on the basis of the control hydraulic pressure _{P SLT} by the line pressure modulator NO.2 valve 128 as an original pressure.

In the manual valve 120, the output oil pressure _PLM2 is supplied to the input port 120a. When the shift lever 74 is operated to the “D” position or the “L” position, the output hydraulic pressure _PLM2 is supplied to the forward clutch C1 via the forward output port 120f as the forward travel output pressure and the reverse brake. The oil passage of the manual valve 120 is switched so that the hydraulic oil in B1 is drained (discharged) to the atmospheric pressure, for example, from the reverse output port 120r through the discharge port EX, and the forward clutch C1 is engaged and reversely moved. The brake B1 is released.

Further, when the shift lever 74 is operated to the "R" position, output pressure P _LM2 is the hydraulic fluid in the fed and the forward clutch C1 to the reverse brake B1 via the reverse output port 120r as reverse running output pressure Is switched from the forward output port 120f through the discharge port EX to the atmospheric pressure, for example, so that the oil passage of the manual valve 120 is switched, the reverse brake B1 is engaged, and the forward clutch C1 is released. Be made.

  When the shift lever 74 is operated to the “P” position and the “N” position, the oil path from the input port 120a to the forward output port 120f and the oil path from the input port 120a to the reverse output port 120r are both And the oil passage of the manual valve 120 is switched so that the hydraulic oil in the forward clutch C1 and the reverse brake B1 is drained from the manual valve 120, and both the forward clutch C1 and the reverse brake B1 are connected. Be released.

The speed ratio control valve UP114 is closed the suppliable and output port 114k from the input port 114i to the line pressure P _L by being movable in the axial direction to the input side variable pulley 42 via the input and output ports 114j A spool valve element 114a positioned at an upshift position and an original position where the input-side variable pulley 42 communicates with the input / output port 114k via the input / output port 114j, and the spool valve element 114a toward the original position side. A spring 114b as an urging means for _urging , an oil chamber 114c that accommodates the spring 114b and receives the control hydraulic pressure _PDS2 to _apply a thrust toward the original position to the spool valve element 114a, and the spool valve element 114a control pressure P to apply a thrust force toward the upshift position side in And an oil chamber 114d that accepts _S1.

Further, the transmission ratio control valve DN116 is provided so as to be movable in the axial direction, whereby a downshift position where the input / output port 116j communicates with the discharge port EX and an original position where the input / output port 116j communicates with the input / output port 116k. A spool valve element 116a positioned at the first position, a spring 116b as an urging means for urging the spool valve element 116a toward the original position, and a spring 116b that accommodates the spool valve element 116a in the original position side. An oil chamber 116c that receives the control hydraulic pressure _PDS1 to apply a thrust toward the engine, and an oil chamber 116d that receives the control hydraulic pressure _PDS2 to apply a thrust toward the downshift position to the spool valve element 116a. .

In the transmission ratio control valve UP114 and the transmission ratio control valve DN116 thus configured, in the closed state in which the spool valve element 114a is held in the original position in accordance with the urging force of the spring 114b as shown in the left half of the center line, The input / output port 114j and the input / output port 114k are communicated with each other, and the hydraulic oil in the input side variable pulley 42 (input side hydraulic cylinder 42c) is allowed to flow to the input / output port 116j. In the closed state in which the spool valve element 116a is held in the original position according to the urging force of the spring 116b as shown in the right half of the center line, the input / output port 116j and the input / output port 116k are communicated with each other, and the thrust ratio it is allowed that the thrust ratio control oil pressure P _tau from control valve 118 to flow to the input-output port 114k.

Further, when the control oil pressure _PDS1 is supplied to the oil chamber 114d, the spool valve element 114a is increased against the urging force of the spring 114b by a thrust according to the control oil pressure _PDS1 as shown in the right half of the center line. is moved to the shift position side, the line pressure _{P L} is supplied to the control oil pressure _{P DS1} to the corresponding flow rate at the input port output from 114i port 114j menstrual the input side hydraulic cylinder 42c, input and output ports 114k is blocked Accordingly, the flow of the hydraulic oil to the speed ratio control valve DN116 side is prevented. As a result, the shift control pressure Pin is increased, the V groove width of the input side variable pulley 42 is narrowed, and the speed ratio γ is decreased, that is, the continuously variable transmission 18 is upshifted.

When the control oil pressure _PDS2 is supplied to the oil chamber 116d, the spool valve element 116a is lowered against the urging force of the spring 116b by the thrust according to the control oil pressure _PDS2 , as shown in the left half of the center line. The hydraulic fluid in the input side hydraulic cylinder 42c is discharged from the input / output port 114j through the input / output port 114k and the input / output port 116j through the input / output port 116j at a flow rate corresponding to the control hydraulic pressure _PDS2 . As a result, the shift control pressure Pin is lowered, the V-groove width of the input side variable pulley 42 is increased, and the speed ratio γ is increased, that is, the continuously variable transmission 18 is downshifted.

Thus, the line pressure P _L is a used as the basic pressure of the shift control pressure Pin, the control pressure P _DS1 is to be output speed ratio control line pressure P _L input to the valve UP114 input side hydraulic cylinder When the control hydraulic pressure _PS2 is output, the hydraulic oil in the input side hydraulic cylinder 42c is discharged from the discharge port EX and the shift control pressure Pin is reduced. Lowered and continuously downshifted.

For example, as shown in FIG. 4, it is actually determined from a previously stored relationship (shift map) between the vehicle speed V and the target input shaft rotational speed N _IN ^* that is the target input rotational speed of the continuously variable transmission 18 using the accelerator opening Acc as a parameter. The target input shaft rotational speed N _IN ^* set based on the vehicle state indicated by the vehicle speed V and the accelerator opening Acc matches the actual input shaft rotational speed (hereinafter referred to as the actual input shaft rotational speed) N _IN. As described above, the speed change of the continuously variable transmission 18 is executed by feedback control, that is, the V groove widths of the variable pulleys 42 and 46 are changed by supplying and discharging the hydraulic oil to and from the input side hydraulic cylinder 42c, thereby changing the gear ratio γ. Is continuously changed by feedback control.

The shift map in FIG. 4 corresponds to the shift conditions, and the target input shaft rotational speed N _IN ^* is set such that the larger the vehicle speed V is and the larger the accelerator opening Acc is, the larger the gear ratio γ is. Further, since the vehicle speed V corresponds to the output shaft rotation speed _{N OUT,} which is the target value of the input shaft rotational speed _{N IN} target input shaft rotational speed _{N IN} ^* is the target speed ratio ^{_{γ * (= N IN * /}} N OUT) Is determined within the range of the minimum speed ratio γmin and the maximum speed ratio γmax of the continuously variable transmission 18.

Further, the control hydraulic pressure _PDS1 is supplied to the oil chamber 116c of the transmission ratio control valve DN116, and regardless of the control hydraulic pressure _PDS2 , the transmission ratio control valve DN116 is closed to limit the downshift, while the control hydraulic pressure _{PDS2 changes} the speed. The oil ratio is supplied to the oil chamber 114c of the ratio control valve UP114, and regardless of the control oil pressure _PDS1 , the transmission ratio control valve UP114 is closed to prohibit the upshift. That is, the control when the hydraulic _{P DS1} and the control pressure _{P DS2} are not supplied together but of course, also, the speed change ratio control valve UP114 and speed ratio control valve when the control oil pressure _{P DS1} and the control pressure _{P DS2} is supplied together Each of the DNs 116 is in a closed state held in its original position. As a result, one of the solenoid valves DS1 and DS2 does not function due to a failure in the electrical system, and a sudden upshift occurs even when the control hydraulic pressure _PDS1 or the control hydraulic pressure _PDS2 continues to be output at the maximum pressure. It is possible to prevent a downshift or a belt slip due to the sudden shift.

The clamping force control valve 110, via an output port 110t to line pressure P _L by opening and closing an input port 110i from the input port 110i output side variable pulley 46 and the thrust ratio control by being movable in the axial direction valve 118 The spool valve element 110a that enables the belt clamping pressure Pd to be supplied, the spring 110b as an urging means that urges the spool valve element 110a in the valve opening direction, and the spring 110b is accommodated in the spool valve element 110a. An oil chamber 110c that receives the control hydraulic pressure _PSLS to give thrust in the valve opening direction, and feedback that receives belt clamping pressure Pd output from the output port 110t to give thrust in the valve closing direction to the spool valve element 110a. The thrust in the valve closing direction is applied to the oil chamber 110d and the spool valve element 110a. And an oil chamber 110e that accepts modulator pressure P _M in order to impart.

In the clamping pressure control valve 110 thus configured, by the transmission belt 48 is line pressure P _L is continuously regulated pressure control control oil pressure P _SLS so as not slip as a pilot pressure, an output port 110t From this, the belt clamping pressure Pd is output. Thus, the line pressure P _L is used as the basic pressure of the belt clamping pressure Pd. A hydraulic pressure sensor 130 is provided in the oil passage between the output port 110t and the output side hydraulic cylinder 46c, and the belt clamping pressure Pd is detected by the hydraulic pressure sensor 130.

For example, as shown in FIG. 5, a relationship (belt) that is experimentally obtained in advance and stored so that belt slip does not occur between the transmission gear ratio γ and the belt clamping pressure Pd ^* using the accelerator opening Acc corresponding to the transmission torque as a parameter. The belt clamping pressure Pd of the output side hydraulic cylinder 46c is obtained so that the belt clamping pressure Pd ^* determined (calculated) based on the vehicle state indicated by the actual gear ratio γ and the accelerator opening Acc is obtained from the clamping pressure map). The pressure is regulated, and the belt clamping pressure Pd ^*, that is, the frictional force between the variable pulleys 42 and 46 and the transmission belt 48 is increased or decreased according to the belt clamping pressure Pd.

The thrust ratio control valve 118, a thrust ratio control oil pressure P the line pressure _{P L} by opening and closing an input port 118i by being movable in the axial direction from the input port 118i via an output port 118t to the speed ratio control valve DN116 a spool valve element 118a that can supply _τ , a spring 118b as an urging means that urges the spool valve element 118a in the valve opening direction, and the spring 118b is accommodated in the valve opening direction in the valve opening direction. an oil chamber 118c that receives the belt clamping pressure Pd to apply a thrust force, a feedback oil chamber for receiving the thrust ratio control oil pressure P _tau output from the output port 118t to apply a thrust force in the valve closing direction to the spool valve element 118a 118d.

In the thrust ratio control valve 118 configured as described above, the pressure receiving area of the belt clamping pressure Pd in the oil chamber 118c a, the pressure receiving area of the thrust ratio control oil pressure P _tau in the feedback oil chamber 118d b, the biasing force of the spring 118b When F _S, an equilibrium state in the following equation (3).
_Pτ × b = Pd × a + F _S (3)
Therefore, the thrust ratio control hydraulic pressure _Pτ is expressed by the following equation (4) and is proportional to the belt clamping pressure Pd.
_Pτ = Pd × (a / b) + F _S / b (4)

Then, the control oil pressure _P or _DS1 and the control pressure _{P DS2} is not supplied together, or the predetermined pressure or more control pressure _{P DS1} and the predetermined pressure or more control pressure _{P DS2} is both supplied, the speed ratio control valve UP114 and the speed ratio control when the valve DN116 has both been a closed state is held in the original position, since the thrust ratio control oil pressure P _tau is supplied to the input side hydraulic cylinder 42c, and a shift control pressure Pin is the thrust ratio control oil pressure P _tau Matched. That is, the thrust ratio control oil pressure P _tau i.e. the shift control pressure Pin maintain a predetermined relationship between the ratio between the shift control pressure Pin and the belt clamping pressure Pd by the thrust ratio control valve 118 is output.

For example, because of the accuracy of the vehicle speed sensor 58, the detection accuracy of the vehicle speed V is inferior at an extremely low vehicle speed less than a predetermined vehicle speed. Therefore, the rotational speed difference (deviation) ΔN _IN (= N _IN ^* In place of the feedback control of the gear ratio γ for eliminating -N _IN ), the control hydraulic pressure _PDS1 and the control hydraulic pressure _PDS2 are not supplied, and the transmission ratio control valve UP114 and the transmission ratio control valve DN116 are both closed. So-called closing control is executed. Thus, when starting the vehicle, Pin proportional to Pd is supplied to the input side hydraulic cylinder 42c so that the ratio between the transmission control pressure Pin and the belt clamping pressure Pd is set in a predetermined relationship. The belt slip of the transmission belt 48 at the vehicle speed is prevented, and at this time, for example, the thrust ratio τ corresponding to the maximum gear ratio γmax (= output-side hydraulic cylinder thrust W _OUT / input-side hydraulic cylinder thrust W _IN ; W _OUT is the belt sectional area of the clamping Pd × output side hydraulic cylinder 46c, W _iN is the first term on the right side of the shift control pressure Pin × input side hydraulic cylinder 42c cross-sectional area) larger thrust ratio τ capable as above formula (4) When (a / b) and FS / b are set, a good start is performed at the maximum gear ratio γmax or a gear ratio γmax ′ in the vicinity thereof. The predetermined vehicle speed is a lower limit vehicle speed at which a predetermined feedback control can be executed as a vehicle speed V at which the rotational speed of the predetermined rotating member, for example, the input shaft rotational speed _NIN cannot be detected. It is set at about 2km / h.

FIG. 6 is a functional block diagram for explaining the main part of the control function by the electronic control unit 50. In FIG. 6, the target input rotation setting means 150 determines the input shaft rotation speed N _IN based on the vehicle state indicated by the actual vehicle speed V and the accelerator opening Acc from a shift map stored in advance as shown in FIG. Set the target input shaft speed N _IN ^* .

The shift control means 152 is arranged so that the actual input shaft rotational speed N _IN matches the target input shaft rotational speed N _IN ^* set by the target input rotational setting means 150, in other words, the rotational speed difference ΔN _IN (= N _IN ^* -N _IN ) is executed so that the shift of the continuously variable transmission 18 is feedback-executed. That is, the input side hydraulic shift control command signal for changing the V groove widths of both variable pulleys 42 and 46 by controlling the flow of hydraulic fluid to the cylinder 42c (hydraulic pressure command) is output to S _T to the hydraulic control circuit 100 shift The ratio γ is continuously changed.

The belt clamping pressure setting means 154, for example, from the belt clamping pressure map obtained and stored experimentally in advance as shown in FIG. 5, for example, the actual accelerator opening Acc and the actual input shaft rotational speed N by the electronic control unit 50. _The belt clamping pressure Pd ^* is set based on the vehicle state indicated by the actual speed ratio γ (= N _IN / N _OUT ) calculated based on _IN and the output shaft rotational speed N _OUT . That is, the belt clamping pressure setting means 154 sets the belt clamping pressure Pd of the output side hydraulic cylinder 46c for obtaining the belt clamping pressure Pd ^* .

Belt clamping pressure control means 156, the belt clamping pressure setting means 154 clamping pressure control command signal S _B for pressurizing regulating the belt clamping pressure Pd of the output side hydraulic cylinder 46c for the set belt clamping pressure Pd ^* is obtained by Output to the hydraulic control circuit 100 to increase or decrease the belt clamping pressure Pd ^* .

The hydraulic control circuit 100, the supply of hydraulic fluid by operating the solenoid valve DS1 and the solenoid valve DS2 so shifting of the continuously variable transmission 18 is executed to the input side hydraulic cylinder 42c in accordance with the shift control command signal S _T · to control the emissions, by operating the linear solenoid valve SLS so that the belt clamping pressure Pd ^* is increased or decreased pressure of the belt clamping pressure Pd adjusted in accordance with the above clamping force control command signal S _B.

The engine output control means 158 outputs an engine output control command signal S _E , for example, a throttle signal, an injection signal, an ignition timing signal, etc., to the throttle actuator 76, the fuel injection device 78, and the ignition device 80, respectively, for output control of the engine 12. To do. For example, the engine output control means 158 controls the engine torque T _E and outputs a throttle signal for opening and closing the electronic throttle valve 30 such that the throttle opening theta _TH corresponding to the accelerator opening Acc to the throttle actuator 76.

By the way, in order to generate a shift control pressure Pin (hereinafter referred to as a required Pin pressure) and a belt clamping pressure Pd ^* that are necessary when the speed ratio γ is feedback controlled in order to improve the speed change performance, for example, the speed change response. the belt clamping pressure Pd (hereinafter, Pd referred pressure) to protect itself, it is necessary to set the line pressure P _L as the original pressure thereof needed Pin pressure and the Pd pressure.

In other words, the line pressure P _L does not readily occur even good belt slippage relatively high shifting response than the needed Pin pressure and the Pd pressure, but becomes a factor of fuel efficiency is deteriorated unnecessarily high. Also, or a is lower than the needed Pin pressure line oil pressure P _L a factor shift response is reduced, the line pressure P _L is easily made factors occur less belt slippage than the Pd pressure.

Accordingly, the line hydraulic pressure setting means 160 is controlled by the linear solenoid valve SLT independently of the flow rate control of the hydraulic oil to the input side hydraulic cylinder 42c by the solenoid valves DS1 and DS2 and the control of the Pd pressure by the linear solenoid valve SLS. that the line pressure P _L, is set based on either the higher oil pressure of the needed Pin pressure and the Pd pressure.

For example, the line oil pressure setting means 160, on the basis of the needed Pin pressure and respectively the line pressure P _L which is the higher hydraulic pressure of the hydraulic pressure by adding a predetermined margin value to the control accuracy and vehicle condition considering of the Pd pressure to set the line hydraulic pressure _{P L.}

Figure 7 is a diagram for explaining an example of a concept of setting the line hydraulic pressure P _L. As shown in FIG. 7, the reference margin value EX determined experimentally in advance as the predetermined margin value is the required Pin pressure reference margin value EXin at the required Pin pressure, and the Pd pressure reference margin at the Pd pressure. The value EXd. Thus, different values are set for the reference margin value EX for the required Pin pressure and the Pd pressure. Then, it was added needed Pin pressure reference margin value EXin to the needed Pin pressure hydraulics and any higher hydraulic pressure is required line pressure P _L or target line of hydraulic and which was added Pd pressure reference margin value EXd to the Pd pressure It is set as a hydraulic pressure (hereinafter, target line hydraulic pressure) P _L ^* .

The setting of the target line oil pressure P _L ^* will be described in detail below.

Pd pressure is directly regulated pressure control by operating the linear solenoid valve SLS accordance clamping pressure control command signal S _B in accordance with the belt clamping pressure control means 156 such that the Pd pressure for the belt clamping pressure Pd ^* is obtained It is hydraulic. Therefore, as the Pd pressure used for setting the target line oil pressure P _L ^* , the Pd pressure for obtaining the belt clamping pressure Pd ^* set by the belt clamping pressure setting means 154 is used as it is.

On the other hand, the required Pin pressure is a hydraulic pressure generated as a result of the flow control of the hydraulic oil and the belt clamping pressure control when feedback control is performed so that the target input shaft rotational speed N _IN ^* (or the target speed ratio γ ^* ) is obtained. However, pressure regulation control is not directly performed. Therefore, the necessary Pin pressure used for setting the target line oil pressure P _L ^* is calculated as an estimated value.

The line oil pressure setting means 160 is previously experimentally used to estimate the necessary Pin pressure using the speed ratio γ, the change speed d (γ ^* ) / dt of the target speed ratio γ ^* , the input torque T _IN , the Pd pressure, and the like as variables. From the relationship obtained and stored (calculation formula, required Pin pressure = f (γ, d (γ ^* ) / dt, T _IN , Pd pressure)), the actual gear ratio γ and the target input rotation setting means 150 The change speed d (γ ^* ) / dt of the target speed ratio γ ^* (= N _IN ^* / N _OUT ) calculated by the electronic control unit 50 based on the set target input shaft rotational speed N _IN ^* , the input torque T _The necessary Pin pressure is calculated based on the estimated engine torque value T _E0 corresponding to _IN , the Pd pressure for obtaining the belt clamping pressure Pd ^* set by the belt clamping pressure setting means 154, and the like.

Engine torque calculating means 162, for example, experimentally determined in advance and stored relationship between the engine rotational speed N _E and the engine torque estimated value T _E0 throttle valve opening theta _TH as shown in FIG. 8 as a parameter (engine torque From the map), an estimated engine torque value T _E0 estimated based on the actual engine speed _NE and the throttle valve opening θ _TH is calculated.

As described above, the Pd pressure is a hydraulic pressure that is directly pressure-controlled so as to be a Pd pressure for obtaining the belt clamping pressure Pd ^* . Thus, the line oil pressure setting means 160, the only consideration of variations of the actual line oil pressure P _L to the line oil pressure control command signal S _PL, i.e. the actual line oil pressure P _L to the line pressure control command signal S _PL its actual line pressure P _L even variations in to exceed the Pd pressure, set the Pd pressure reference margin value EXd stored predetermined experimentally.

On the other hand, as described above, the necessary Pin pressure is not a pressure regulation control directly, but a hydraulic pressure calculated as an estimated value. In addition, a certain margin is required to perform feedback control so as to maintain a constant gear ratio γ. Thus, the line oil pressure setting means 160, the line oil pressure control command signal S _PL actual line pressure P _L of only to be larger than the Pd pressure reference margin value EXd that has been set in consideration variation and, in advance experimentally The required Pin pressure reference margin value EXin determined and stored in (1) is set. In this way, different reference margin values EX are set for the required Pin pressure and the Pd pressure.

  Further, the line hydraulic pressure setting means 160 may change the Pd pressure reference margin value EXd and the required Pin pressure reference margin value EXin, respectively, based on the vehicle state (running state).

For example, the line oil pressure setting means 160 is configured to change the actual line oil pressure P _L with respect to the line oil pressure control command signal S _PL when the hydraulic oil temperature is low and when the hydraulic oil temperature is low compared to the normal oil temperature when the hydraulic oil temperature is normal. Since the control accuracy decreases, a larger Pd pressure reference margin value EXd and a required Pin pressure reference margin value EXin are set as compared with the normal oil temperature. The normal oil temperature is assumed to be, for example, the oil temperature after completion of warm-up.

  Further, the line hydraulic pressure setting means 160 changes the speed ratio γ during steady running so that a constant speed ratio γ is maintained, and the change in the speed ratio γ is extremely small compared to the speed change that changes the speed ratio γ. Therefore, the Pd pressure reference margin value EXd and the required Pin pressure reference margin value EXin, which are smaller than those at the time of shifting, are set.

Further, the line hydraulic pressure setting means 160 greatly changes the gear ratio γ during traveling that requires a sudden gear change, that is, during a sudden gear change in which the gear ratio γ exceeds a predetermined value and the gear ratio γ is changed. Therefore, a large Pd pressure reference margin value EXd and a necessary Pin pressure reference margin value EXin are set as compared with those during normal shifting and steady running. This normal shift is a shift in which the speed ratio γ is changed within a range in which the speed of change of the speed ratio γ does not exceed a predetermined value. During a sudden shift, the accelerator pedal 68 suddenly returns as indicated by an arrow A in FIG. When the vehicle is operated to perform an off-up shift, which is a rapid acceleration, or when the speed ratio γ is changed along the lower limit value of the target input shaft rotational speed N _IN ^* as shown by the arrow B, the rapid acceleration It is assumed that the vehicle will be running.

The line hydraulic pressure setting means 160 calculates a hydraulic pressure value obtained by adding the Pd pressure reference margin value EXd to the Pd pressure set by the belt clamping pressure setting means 154 as the Pd pressure line hydraulic pressure P _L d, and hydraulic value obtained by adding the needed Pin pressure reference margin value EXin to the Pin pressure calculated as the Pin pressure-purpose line oil pressure P _L in, whichever is higher between the Pd pressure-purpose line oil pressure P _{L d} and the Pin pressure-purpose line oil pressure P _L in Is selected as the target line oil pressure P _L ^* , and the selected higher line oil pressure is set as the target line oil pressure P _L ^* . As shown in FIG. 7, the target line oil pressure P _L ^* is set with the line oil pressure MAX guard that has been experimentally determined in advance so that the line oil pressure P _L does not exceed the allowable load of the transmission belt 48 as an upper limit. The This line oil pressure MAX guard may be determined in consideration of fuel consumption instead of or in addition to the allowable load of the transmission belt 48.

Line pressure control means 164, for example, pre-stored relationship shown in FIG. 9 and the line oil pressure control command signal S _PL to the line pressure P _L which is pressure regulated on the basis of the line pressure control command signal S _PL (line Line pressure control command signal S _PL (control current) for obtaining the target line oil pressure P _L ^* based on the target line oil pressure P _L ^* set by the line oil pressure setting means 160 from the hydraulic pressure characteristic). output to 100 to pressure the line hydraulic pressure _{P L} tone is.

The hydraulic control circuit 100 actuates the linear solenoid valve SLT pressure to line pressure P _L regulated in accordance with the above line oil pressure control command signal S _PL.

Line pressure characteristic of FIG. 9, the line hydraulic pressure control command signal S _PL and the line oil pressure control command signal S relationship between the control hydraulic pressure P _SLT is the output oil pressure of the linear solenoid valve SLT normally open type which is driven based on the _PL Table in relation to the relationship between the line oil pressure _{P L} to be pressure regulated by the primary regulator valve 122, and line hydraulic pressure control command signal _{S PL} and the line pressure _{P L} based, and control pressure _{P SLT} and its control oil pressure _{P SLT} This is a standard hydraulic characteristic obtained experimentally in advance.

By the way, the line hydraulic pressure characteristics as shown in FIG. 9 vary or change due to individual differences of the linear solenoid valve SLT and the primary regulator valve 122 or changes with time. Therefore, the target line pressure P _L ^* a line oil pressure control command signal S actual line hydraulic pressure regulated on the basis of the _PL (hereinafter, actual referred line pressure) there is a possibility that the difference between the higher tolerance occurs and P _L . So as to eliminate the difference between the target line pressure P _L ^* and the actual line oil pressure P _L, we are conceivable to learn correcting the line pressure characteristic on the basis of the target line pressure P _L ^* and the actual line oil pressure P _L .

However, the actual line pressure P _L, as shown in FIG. 10, a hydraulic rise that varies by a change in the engine rotational speed N _E as influencing factors rotational speed or the engine rotational speed N _E of the oil pump 28 Rising includes the amount, since it is the also the engine rotational speed N _E a line oil pressure control command signal S _PL are the same resulting in different actual line pressure P _L, the line pressure without considering the rising amount If the characteristics are learned and corrected, accurate line hydraulic characteristics may not be obtained. That is, there arises a problem that the learning accuracy of the line hydraulic pressure characteristics deteriorates.

Figure 10 is a diagram showing the valve characteristics of the actual line oil pressure P _L with respect to the engine rotational speed N _E of the primary regulator valve 122 corresponding to each target line pressure P _L ^*. 10, each ● point, hydraulic pressure generated from the oil pump 28 is caused to increase with increase in the engine rotational speed N _E, the line oil pressure control command signal S line pressure P to be pressure regulated on the basis of the _PL is the point which reaches the _L, and the boundary point where the target line pressure P _L ^* is obtained at the high speed side of the engine rotational speed N _E to the point as a boundary (hereinafter, referred to as the boundary point between the cracking point) . The rising amount is a hydraulic pressure increase based on the cracking point. Further, the line pressure characteristics than is being learning correction as the target line pressure P _L ^* and the actual line oil pressure P _L is fit at the cracking point. FIG 10 is a characteristic diagram of the actual line oil pressure P _L after being previously learned correction.

11, among the previously experimentally sought stored relationship rising amount is determined and a line pressure P _L and the engine rotational speed N _E as a variable (rising amount map), a typical line pressure P _L and is a table showing the rising amount of the engine rotational speed N _E. As shown in FIG. 11 and FIG. 10, rising amount higher actual line pressure P _L, also tends to increase the higher the engine rotational speed N _E. Difference in this way, since also target line pressure P _L ^* is the same, depending on the engine rotational speed N _E becomes different the actual line oil pressure P _L, the target line pressure P _L ^* and the actual line oil pressure P _L If the line hydraulic pressure characteristic is uniformly corrected based only on this, there is a possibility that an accurate line hydraulic pressure characteristic cannot be obtained.

Accordingly, the learning correction unit 166 learns and corrects the line hydraulic pressure characteristics based on the difference between the target line hydraulic pressure P _L ^* and the actual line hydraulic pressure P _L, and further learns and corrects the line hydraulic pressure characteristics based on the rising amount. To do. Note that the feedback control of the line oil pressure control command signal _SPL may be considered so as to eliminate the difference between the target line oil pressure P _L ^* and the actual line oil pressure P _L. The characteristic learning correction is more preferable.

  The learning correction of the line hydraulic characteristic will be described in detail below.

Actual line oil pressure detecting means 168 detects the actual line oil pressure P _L whose pressure regulated by the linear solenoid valve SLT and the primary regulator valve 122 based on the line hydraulic pressure control command signal S _PL.

In the hydraulic control circuit 100 of this embodiment, since the oil pressure sensor for directly detecting the actual line pressure P _L is not provided, with the hydraulic pressure sensor 130 for detecting the belt clamping pressure Pd, the belt clamping pressure detecting the actual line oil pressure P _L which is the original pressure pd. In other words, the configuration of the clamping force control valve 110, the maximum value of the belt clamping pressure Pd is the actual line oil pressure P _L, the belt clamping pressure Pd is controlled clamping force control valve 110 such that the maximum value, when the detecting the belt clamping pressure Pd as the actual line oil pressure P _L.

More specifically, the actual line hydraulic pressure detecting means 168, the line pressure P _L is output an instruction to the belt clamping pressure control means 156 so that the control oil pressure P _SLS to fully open the input port 110i to the input is output while, the oil pressure sensor 130 when the line hydraulic pressure detection command signal S _B for the control oil pressure P _SLS to fully open the input port 110i by the belt clamping pressure control unit 156 is output _'is output to the hydraulic control circuit 100 detecting the belt clamping pressure Pd that is detected as the actual line hydraulic pressure P _L by.

Hydraulic rise calculating means 170, the target line pressure P from the rising amount map Rising amount is determined and a line pressure P _L and the engine rotational speed N _E as illustrated as a variable in Figure 11 _L ^* and the engine speed N _E The rising amount is calculated based on

The learning correction means 166 includes a hydraulic learned value calculating means 172 for calculating the hydraulic pressure learning value in the target line pressure P _L ^* line oil pressure control command signal corresponding to S _PL, currently stored on the basis of the hydraulic pressure learning value The line hydraulic characteristics are updated as needed to correct the learning.

The hydraulic learning value calculation section 172, the target and subtracting the actual line oil pressure P _L detected by the actual the actual line oil pressure detecting means 168 from the target line pressure P _L ^* set by the line oil pressure setting means 160 to calculate the hydraulic pressure learning value at the line pressure P _L ^* line oil pressure control command signal corresponding to S _PL, corresponding to that the from the hydraulic learned value hydraulic rise its target line pressure calculated by the calculation means 170 P _L ^* The final hydraulic pressure learning value is calculated by subtracting the rising amount.

The learning correction means 166 calculates the oil pressure learning calculated by the oil pressure learning value calculation means 172 from the line oil pressure P _L in the line oil pressure control command signal S _PL corresponding to the target line oil pressure P _L ^* in the currently stored line oil pressure characteristics. the subtracted value a value, updated at any time as the line pressure P _L after correction in the line oil pressure control command signal S _PL. The line pressure P _L to be updated may be a continuous characteristic by linear interpolation or the like because it is discrete.

  FIG. 12 is a flowchart for explaining a main part of the control operation of the electronic control unit 50, that is, a control operation for accurately learning and correcting the line hydraulic pressure characteristic. Is to be executed.

First, in a step (hereinafter, step is omitted) S1 corresponding to the line hydraulic pressure setting means 160, the hydraulic pressure value obtained by adding the Pd pressure reference margin value EXd to the Pd pressure set from the belt clamping pressure map is a Pd pressure line. Calculated as the oil pressure P _L d and the necessary Pin pressure calculated from the relationship obtained experimentally and stored in advance (required Pin pressure = f (γ, d (γ ^* ) / dt, T _IN , Pd pressure)) subject to the needed Pin pressure reference margin value EXin hydraulic pressure value is calculated as a Pin pressure-purpose line oil pressure P _L in, whichever is higher line oil pressure of this Pd pressure-purpose line oil pressure P _{L d} and the Pin pressure-purpose line oil pressure P _L in Is selected, and the selected higher line oil pressure is set as the target line oil pressure P _L ^* .

Then, the step S2 corresponding to the actual line oil pressure detecting means 168, the actual line pressure P _L whose pressure regulated by the linear solenoid valve SLT and the primary regulator valve 122 based on the line hydraulic pressure control command signal S _PL is detected. For example, the command control pressure P _SLS to fully open the input port 110i to the line pressure P _L is input is output is output, for controlling hydraulic pressure P _SLS to the input port 110i is fully opened is outputted line pressure detection command signal S _{B 'of} the belt clamping pressure Pd is to be detected is detected as the actual line hydraulic pressure P _L by the oil pressure sensor 130 when it is output to the hydraulic control circuit 100.

Then, the in S3 corresponding to the hydraulic pressure rise calculating means 170, the line pressure as illustrated in FIG. 11 P _L and the engine rotational speed N _E and the target line pressure from the rising amount map Rising amount is defined as the variable P _L ^* and the amount of writhing is calculated based on the engine rotational speed N _E.

Next, in S4 corresponding to the oil pressure learned value calculation means 172, the actual line oil pressure P _L detected in S2 is subtracted from the target line oil pressure P _L ^* set in S1, and the target line oil pressure P _L ^* hydraulic learned value is calculated in the line oil pressure control command signal S _PL corresponding to, is further rising amount subtraction corresponding to the hydraulic that goal line is calculated from the learned value at the S3 hydraulic P _L ^* final A hydraulic pressure learning value is calculated.

Then, in the above learning correction means 166 S5 corresponding to the current stored target line pressure P is set in line pressure characteristics are in the S1 _L ^* line oil pressure control command signal corresponding to the S line in the _PL pressure P _L values hydraulic learned value calculated by the S4 is subtracted from, is updated at any time as the line pressure P _L after correction in the line oil pressure control command signal S _PL. The line pressure P _L is the update is a continuous characteristic by linear interpolation or the like because it is discrete. In this way, the currently stored line hydraulic pressure characteristic is updated as needed to correct learning.

Here, in the case of using the actual line pressure P _L in the control of the continuously variable transmission 18, as described above, the detection value by consideration of the hydraulic pressure sensor hydraulic response delay preferably Suitable Although feedback control to the learning control It is not considered. Then, it is necessary to calculate the actual line pressure P _L as the estimation value when the feedback control as the speed change control is performed. Of course, it is required that the estimated value be obtained with high accuracy. For example, when the actual line pressure P _L and the required calculate the flow rate of the hydraulic oil based on the differential pressure between the pressure Pin to the input side variable pulley 42 (the input side hydraulic cylinder 42c) to such performs a shift control, the actual line estimate of hydraulic P _L is required to be determined accurately.

Therefore, line pressure estimated value calculating means 174, the learning correction means 166 by learning corrected line pressure characteristic from the line oil pressure control unit 164 is output and line pressure controlled by the command signal S _PL actual line pressure P based on to calculate the estimated value of _L, it is corrected on the basis of the rising amount corresponding to the target line pressure P _L ^* of this time which is calculated an estimate of the actual line oil pressure P _L by the oil pressure rise calculating means 170. This correction, the estimated value of the actual line oil pressure P _L is intended to adding the rising amount, thereby calculating the final estimate of the actual line oil pressure P _L.

Figure 13 is a flowchart illustrating a control operation for calculating the estimated value accurately of a substantial part i.e. the actual line hydraulic pressure P _L of the control operation of the electronic control unit 50, a very short of the order of several msec to several tens msec e.g. It is executed repeatedly at cycle time.

First, the in-line estimated hydraulic pressure calculation means 174 S11 corresponding to the Figure 12 S1 to S5 in learning corrected actual line oil pressure control command from the line hydraulic pressure characteristic signal S _PL actual line pressure P _L based on the Is estimated.

Then, the step S12 which corresponds to the hydraulic pressure rise calculating means 170, the line pressure as illustrated in FIG. 11 P _L and the engine rotational speed N _E and the target line pressure from the rising amount map Rising amount is defined as the variable P _L ^* and the amount of writhing is calculated based on the engine rotational speed N _E.

Then, the line in S13 described corresponding to the estimated hydraulic pressure calculation means 174, the S11 rising amount calculated by the S12 in the estimate of the actual line oil pressure P _L which is calculated is added by finally real line estimate of hydraulic P _L is calculated.

As described above, according to this embodiment, and the rotational speed influencing factors of the oil pump 28 which is calculated by the actual line oil pressure P _L and the hydraulic increment calculation means 170 detected by the actual line oil pressure detecting means 168 Rising based on the amount, the learning correction means 166 by the line oil pressure control command signal S _PL and the line oil pressure control command signal S prestored line pressure characteristic learning correction of the line pressure P _L which is pressure regulated on the basis of the _PL Therefore, an appropriate line hydraulic pressure characteristic can be obtained regardless of the rising amount, and the learning accuracy of the line hydraulic characteristic is improved. Thus, for example, the line oil pressure control command signal _SPL can be set so that the line oil pressure P _L is as small as possible within the range where the target line oil pressure P _L ^* can be obtained. It was compared to when setting the line hydraulic pressure control command signal S _PL to the line pressure P _L higher, thereby improving fuel efficiency loss is suppressed in the oil pump 28.

Further, according to this embodiment, the line pressure estimation value calculating means 174, an estimate of the actual line oil pressure P _L based on the learning correction means 166 by learning corrected line pressure control command from the line hydraulic pressure characteristic signal S _PL together are calculated, the estimated value of the actual line oil pressure P _L is corrected based on the rising amount, the continuously variable transmission 18 based on the estimated value of the corrected real line pressure P _L is controlled by an electronic control unit 50 Runode, estimation accuracy of the actual line oil pressure P _L is increased. Thereby, the control accuracy of the continuously variable transmission 18 is improved. For example, when the flow rate of hydraulic fluid to the input side hydraulic cylinder 42c is calculated based on the differential pressure between the actual line oil pressure P _L and the required Pin pressure and the shift control is performed by feedback control, the actual line oil pressure P _L The estimated value is obtained with high accuracy, and the shift control accuracy is improved.

  As mentioned above, although the Example of this invention was described in detail based on drawing, this invention is applied also in another aspect.

For example, in the illustrated embodiment, the hydraulic control circuit 100, oil pressure sensor for directly detecting the actual line pressure P _L is not provided, the actual line hydraulic pressure detecting means 168, for detecting the belt clamping pressure Pd Although the actual line oil pressure P _L is detected using the oil pressure sensor 130, the actual line oil pressure P _L may be directly detected by providing a hydraulic pressure sensor that detects the actual line oil pressure P _L.

  In the above-described embodiment, the estimated value calculated by the line oil pressure setting unit 160 is used as the necessary Pin pressure. However, a hydraulic pressure sensor for detecting the necessary Pin pressure is provided, and the sensor pressure is replaced with the estimated value. May be used.

Further, the input shaft rotational speed N _IN and the related target input shaft rotational speed N _IN ^* in the above-described embodiment are replaced with the input rotational speed N _{IN and the} like, and the engine rotational speed _NE and the related target. The engine rotational speed N _E ^* or the like, or the turbine rotational speed _NT or a target turbine rotational speed _NT ^* related thereto may be used.

  In the above-described embodiment, the torque converter 14 provided with the lock-up clutch 26 is used as the fluid transmission device. However, the lock-up clutch 26 is not necessarily provided. In addition, other fluid type power transmission devices such as a fluid coupling (fluid coupling) having no torque amplification function may be used.

  The above description is only an embodiment, and the present invention can be implemented in variously modified and improved forms based on the knowledge of those skilled in the art.

1 is a skeleton diagram illustrating a vehicle drive device to which the present invention is applied. It is a block diagram explaining the principal part of the control system provided in the vehicle in order to control the vehicle drive device etc. of FIG. FIG. 3 is a hydraulic circuit diagram showing a main part relating to belt clamping pressure control of a continuously variable transmission, gear ratio control, and engagement hydraulic control of a forward clutch or reverse brake accompanying operation of a shift lever in the hydraulic control circuit. It is a figure which shows an example of the shift map used when calculating | requiring a target input rotational speed in the shift control of a continuously variable transmission. It is a figure which shows an example of the belt clamping pressure map which calculates | requires belt clamping pressure according to gear ratio etc. in clamping pressure control of a continuously variable transmission. It is a functional block diagram explaining the principal part of the control function of the electronic control apparatus of FIG. It is a figure for demonstrating an example of the way of thinking which sets line oil pressure. It is a figure which shows an example of the relationship (engine torque map) calculated | required experimentally beforehand and memorize | stored by using the throttle valve opening as a parameter as an engine rotational speed and an engine torque estimated value. It is a figure which shows an example of the relationship (line oil pressure characteristic) memorize | stored beforehand with a line oil pressure control command signal and the line oil pressure adjusted based on the line oil pressure control command signal. It is a figure which shows an example of the valve characteristic of the actual line oil pressure with respect to the engine speed of a primary regulator valve corresponding to each target line oil pressure. An example of a rising amount at a typical line pressure and engine rotational speed in a relationship (rising amount map) that has been experimentally obtained and stored in advance, in which the rising amount is determined with the line pressure and the engine rotational speed as variables. It is a chart which shows. 3 is a flowchart for explaining a control operation for accurately learning and correcting a main part of the control operation of the electronic control device of FIG. 3 is a flowchart for explaining a control operation for accurately calculating a main part of the control operation of the electronic control device of FIG. 2, that is, an estimated value of an actual line oil pressure.

Explanation of symbols

18: Continuously variable transmission (automatic transmission)
28: Oil pump 42: Input side variable pulley (drive side pulley)
46: Output side variable pulley (driven pulley)
48: Transmission belt (belt)
50: Electronic control device (control device)
122: Primary regulator valve (pressure regulator)
164: Line oil pressure control means 166: Learning correction means 168: Actual line oil pressure detection means 170: Oil pressure increase calculation means 174: Line oil pressure estimated value calculation means SLT: Linear solenoid valve (pressure regulator)

Claims (2)

A pressure regulator that regulates the hydraulic oil discharged from the oil pump to the line hydraulic pressure based on the hydraulic pressure command value, and the necessary line based on the required line hydraulic pressure from the prestored relationship between the hydraulic pressure command value and the line hydraulic pressure A line oil pressure control means for outputting an oil pressure command value for obtaining oil pressure, an actual line oil pressure detection means for detecting an actual line oil pressure adjusted by the pressure adjusting device, and an actual line oil pressure detection means detected by the actual line oil pressure detection means. A control device for an automatic transmission comprising learning correction means for learning and correcting the relationship based on actual line hydraulic pressure,
An oil pressure increase calculating means for calculating the oil pressure increase of the line oil pressure based on the oil pump rotation speed based on the oil pump rotation speed;
The automatic transmission control device, wherein the learning correction means further learns and corrects the relationship based on the hydraulic pressure increase calculated by the hydraulic pressure increase calculation means. The automatic transmission includes a driving pulley, a driven pulley, and a belt wound around both pulleys. A transmission ratio is controlled by hydraulic oil acting on the driving pulley and the driven pulley is connected to the driven pulley. It is a continuously variable transmission in which the belt clamping pressure is controlled by the working hydraulic oil,
A line oil pressure estimated value calculating means for calculating an estimated value of the line oil pressure based on the oil pressure command value from the relationship corrected by the learning correction means and correcting the estimated value of the line oil pressure based on the amount of increase in the oil pressure. Further comprising
2. The automatic transmission control device according to claim 1, wherein the continuously variable transmission is controlled based on the corrected estimated value of the line oil pressure.

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